The invention relates to a two stage compressor for a gas turbine engine with a mixed flow first stage, a centrifugal second stage, and an intermediate diffusing duct.
Although mixed flow, or diagonal flow, compressors are well known to those skilled in the art of gas turbine engine design, the commercial adoption of mixed flow compressors, particularly in aircraft engines, has been very limited. Most aircraft engines to-date utilise axial flow compressor rotors, centrifugal flow compressors or a combination of both.
The configuration and design of axial flow compressors and centrifugal flow compressors is considered very well known in the art and it is only necessary here to present a general outline of their features and advantages or disadvantages.
Large diameter gas turbine engines are generally constructed with multiple exclusively axial stages generally arranged downstream of an intake fan and by-pass duct. In conventional engines the low pressure axial compressor is mounted on the same shaft as the fan and a low pressure turbine, and the high pressure compressor rotors are mounted on a coaxial high pressure shaft driven by a high pressure turbine. Understandably these multiple stage axial compressors are large and complex machines. They can be justified for their high efficiency in high thrust engine constructions.
Smaller engines are sometimes constructed with a centrifugal compressor as the terminal high pressure stage with a series of axial low pressure stages upstream. The centrifugal rotor, together with surrounding diffuser ducts, considerably increases and dictates the maximum diameter of the engine and forward surface area. However, especially in smaller engine designs, the centrifugal compressor provides high efficiency and reduces the axial length of the engine at the expense of an increase in the radial dimension.
A third common compressor structure includes two centrifugal compressor rotors; however, commercial adoption of this compressor design is very limited. The duct work required to convey compressed air from the first stage centrifugal compressor to the second stage centrifugal compressor rotor is very complex, difficult to manufacture accurately and assemble. Multiple centrifugal stages subject the engine to a significant weight and air drag penalty with prohibitively increased diameter. The increased bulk of the engine envelope and loss of compressor efficiency, through the complex ducting required between first and second stages, has severely limited the adoption of a two stage centrifugal compressor.
The common combination of a centrifugal compressor with axial low pressure stage also suffers from several disadvantages that are generally accepted by designers as inevitable. The engine envelope diameter is dictated by the centrifugal compressor and surrounding diffuser. The axial compressor is often constructed of two or more axial rotors with stator blade assemblies between each axial stage. As a result the number of blades and rotors significantly adds cost to the engine and mechanical complexity. In the current economy for gas turbine engines the overall engine price has been dropping relative to inflation, whereas the cost of materials and engine design costs have been rising. For example, the cost of titanium used in the axial compressor blades has tripled in the last ten years. Due to high design costs for such complex machine parts, the expedient of conservative design practices has resulted in heavier, more robust blades to ensure an adequate safety margin. Therefore, although the design and construction of axial compressors is well known, increases in material costs and concern over the high cost of designing these compressors has led to a desire for a less complex and economically efficient compressor design.
In general, the fewer rotor stages and stator stages that are required in a compressor, the better. Multiple stages and highly complex geometries significantly increase the costs of compressors. To-date, however, experiments in adopting diagonal flow or mixed flow compressor blades have been inconclusive. For example there is no production gas turbine engine available with a mixed flow compressor to date, although experimental results are well documented.
It is well recognised that the cost and reliability of modern gas turbine engines is significantly determined by the number of compressor stages, or acceleration/diffusion operations within the compressor section. It is long recognised that reducing the number of compressor stages will have beneficial effect on the cost of this equipment. Although centrifugal compressor stages compared with axial flow compressors offer lower cost and higher static pressure ratio, centrifugal compressors are slightly less efficient and penalise the design with a larger outer engine envelope diameter than a comparable axial flow compressor. The axial flow compressor of course has a longer axial dimension but suffers from a lower resistance to foreign object damage as well as a lower tolerance to distortion and non-uniformity of inlet airflow distribution. On the other hand using multiple centrifugal compressor stages occasions large aerodynamic losses in the duct work required between the stages as well as significant penalties in weight, engine complexity and manufacturing costs.
Mixed flow or diagonal flow, compressor stages have been recognised in the prior art as providing advantages over both the axial flow and centrifugal flow compressors. For example, a mixed flow compressor has a more rugged design which is superior in foreign object damage resistance to an axial flow compressor and the length of blades enable designers to increase the blade width significantly strengthening the mixed flow blades in comparison to axial flow blades. In addition the part speed benefits of a mixed flow compressor with a significant radius change reduces stress, increases part and bearing life when compared with a large diameter centrifugal compressor rotor. The manufacturing of a mixed flow compressor rotor is somewhat simplified in comparison to a centrifugal compressor and the increase in diameter is significantly lessened.
An example of an adoption of a mixed flow compressor rotor is shown in U.S. Pat. No. 4,678,398 to Dodge et al. In this example the mixed flow compressor rotor is positioned as the first stage upstream of a flow splitter by-pass duct and high pressure axial flow compressor. The mixed flow rotor positioned at the engine inlet with relatively rugged blade construction increases the resistance to foreign object damage and fully utilises the centrifugal effect to propel foreign objects radially outwardly through the by-pass duct thereby protecting the high pressure axial compressor sections downstream.
A significant limitation of the Dodge mixed flow compressor however, is the stat ed objectives which inevitably result in transonic/supersonic air flow speeds in the compressor. The design parameters limit the engine envelope to be comparable to that of an axial flow compressor whereas the design objective is to attain the static pressure ratio, cost and inlet resistance to foreign object damage of a centrifugal compressor. In order to obtain these objectives however, Dodge approaches the mixed flow compressor design by requiring transonic velocities and deals with the need to accommodate sonic shock waves within the structure.
Another example of an attempt to replace several axial compressor stages with a single mixed flow compressor stage as a cost reduction is shown in a paper entitled xe2x80x9cMixed-Flow Compressor Stage Design and Test Results with a Pressure-Ratio of 3:1xe2x80x9d Musgrave, D. S. and Plehn, N. J. presented at Gas Turbine Conference and Exhibition, Anaheim, Calif. May 31, to Jun. 4, 1967. In this example the mixed flow compressor stage was designed to be the terminal stage behind an upstream multi-stage axial compressor. The mixed flow compressor stage has an advantage over a conventional centrifugal stage compressor in that the envelope radius is significantly reduced.
A further example of utilisation of a mixed flow compressor is shown in a paper entitled xe2x80x9cDesign and Rotor Performance of a 5:1Mixed-flow Supersonic Compressorxe2x80x9d by Monig, R., Elmdorf, W. and Gallus, H. E. presented at International Gas Turbine and Aeroengine Congress and Exposition, Cologne, Germany Jun. 1-4, 1992. This paper and experimental results deal extensively with the need to stabilise shock waves within the diagonal or mix flow compressor passage. Strong sonic shock waves significantly reduce efficiency to 75%. Although a relatively high 5:1 compression ratio is achieved, this type of design is impractical due to high stresses from the rapid rotor speed and sonic shock waves.
It is an objective of the present invention to provide a sub-sonic mixed flow compressor stage to derive benefits from low part speeds, resulting low stresses, longer part life and foreign object damage resistance of a mixed flow compressor as well as the increased work capacity derived by significant radius changes in comparison to an axial flow compressor.
It is a further object of the invention to replace several upstream low pressure axial stages with a single mixed flow stage thereby reducing the cost of part manufacturing, engine assembly and maintenance of the compressor.
It is a further object of the invention to provide a mixed flow compressor and intermediate duct upstream of a centrifugal compressor to achieve a combined pressure ratio in the order of 10:1 to 13:1.
Further objects of the invention will be apparent from review of the disclosure, drawings and description of the invention below.
The invention provides a two stage compressor for a gas turbine engine with a mixed flow first stage, a centrifugal second stage, and an intermediate duct.
The mixed flow stage has a mixed rotor rotatable about the central compressor axis with a circumferential array of mixed flow blades between the mixed flow hub and an associated mixed flow shroud.
The downstream centrifugal stage has a centrifugal rotor rotatable about the same compressor axis and possibly but not necessarily on the same shaft. The centrifugal rotor has a circumferential array of radially extending centrifugal flow blades between the centrifugal flow hub and an associated centrifugal flow shroud.
An intermediate duct has an inner duct wall defining an axially curvilinear transition surface of revolution between an outlet end of the mixed flow hub and an inlet end of the centrifugal flow hub, and an outer duct wall defining an axially curvilinear transition surface of revolution between an outlet end of the mixed flow shroud and an inlet end of the centrifugal flow shroud. Preferably, the intermediate duct has a median inlet radius greater than an intermediate duct median outlet radius whereby air flow from the mixed rotor is directed radially inwardly and axially rearwardly toward the centrifugal rotor.
The compressor arrangement has several advantages over prior art compressors. A significant advantage of the invention is the replacement of several axial compressor stages with a single mixed flow stage. Significant savings are achieved in reduction of manufacturing costs and maintenance as well as the axial dimension of the engine is reduced. A sub-sonic mixed flow compressor has the advantage of being more efficient than a supersonic compressor stage as described in the prior art.
The potential for rugged blade design of mixed flow compressor reduces material and manufacturing costs over several axial stages, and reduces stress and increases part life in comparison to a centrifugal compressor rotor. Where utilising two centrifugal compressor stages is impractical and inefficient due to the complex duct work and flow losses in the ducts, using a mixed flow compressor stage upstream of a centrifugal compressor stage eliminates complex duct work while maintaining a benefit of additional work in providing a radial flow of air.
The intermediate duct between the first stage mixed flow and second stage centrifugal flow enables the designer to reduce the inlet diameter of the centrifugal stage. As a result of redirecting the airflow radially inwardly during diffusion and reduction of swirl the diameter of the centrifugal impeller can be reduced thereby reducing stresses in the material of the impeller. The intervening duct may include a single row of stator blades thereby reducing complexity and manufacturing costs, but the scope of the invention is not necessarily limited to single row of stators.
Significantly the shock losses and low efficiency of transonic compressors of the prior art have been overcome by ensuring that the outlet flow from the mixed flow compressor has a subsonic absolute velocity and in addition the intervening duct reduces swirl and diffuses the exit airflow from the mixed stage for introduction at a low relative velocity to the centrifugal compressor inlet. None of the prior art consider this combination of a subsonic mixed flow compressor upstream of a centrifugal compressor. In particular, the prior art does not contemplate severely limiting the compression ratio of a mixed flow first stage in order to combine the mixed flow compressor stage with a terminal centrifugal compressor stage.